Plate-fin heat exchanger



l3,:-izes,496

sept. 15, 1970 L. C. KUN

PLATE-FIN HEAT EXCHANGER Filed Nov. 5,*.1967

3 Sheets-Sheet 1 Wulf "W limp '2 FIGL INVENTOR LESLIE C. KUN BY ATTORNEY Sept. 15,' 1970 Filed Nov. 5, 1967 L. c. KUN

PLATE-fFIN HEAT EXCHANGER 5 Sheets-Sheet 2 HEA'L TRANSFER EFFECTI FINS PER INCH FIGS.

INVENTOR I LEsLlE c, Kun

',n c im ATTORNEY Sept. 15, 1970 L. c. KUN 3,523495 K I INVENTR LESLIE c. KUN

v ATTORNEY United States Patent O 3,528,496 PLATE-FIN HEAT EXCHANGER Leslie C. Kun, Williamsville, N.Y., assignor to Union Carbide Corporation, a corporation of New Yorlr Filed Nov. 3, 1967, Ser. No. 680,498 Int. Cl. F28b 3/00 U.S. Cl. 165-166 13 Claims ABSTRACT F THE DISCLOSURE Flat plates of low thermal conductivity metal are spaced in superimposed relation by short flat fins of high thermal conductivity having opposite edges bonded to the plates, the fins being longitudinally positioned in spaced relation to form first and second fluid passageways with adjacent fins spaced at density of 30-80 per inch in the transverse direction and longitudinally separated by gaps between 0.03 and 0.25 inchlong, with the second fluid passageway gaps superimposed over the first fluid passageway gaps.

BACKGROUND OF THE INVENTION This invention relates to the art of plate-lin type heat exchangers, and particularly to compact constructions having high heat transfer effectiveness.

Heat exchangers of the plate and fin type heretofore used for gas and liquid processing have employed up to about 25 fins/ inch of either the plain or serrated type, with the exchanger parts attached together by brazing, e.g., in a furnace or by dipping into a hot salt bath. To obtain rcasona-ble heat transfer effectiveness, such heat exchangers usually operate in the turbulent flow region, i.e. with Reynolds numbers above about 2000. For many requirements these units produce satisfactory heat transfer coefficients and reasonably good thermal performance. However better performance requires higher flow velocities and/or use of turbulence promoters in the passages, with consequent undesirable increases in pressure drop.

The principal object of this invention is to provide a plate-lin type heat exchanger characterized by higher thermal performance than heretofore attainable, and without higher pressure drop. Another object is to provide a plate-1in type heat exchanger having smaller overall size for a particular heat transfer requirement.

Still another object is to provide an improved method for fabricating a plate-fin type heat exchanger.

These and other objects and novel features will become apparent from the following description and accompanying drawings.

SUMMARY According to this invention, a plate-fin type heat exchanger is provided with at least three flat plates in spaced superimposed relation with the top plate over the intermediate plate and the latter over the bottom plate, each being formed of low thermal conductivity metal. A multiplicity of ilat fins formed of metal having high thermal conductivity at least 4 times that of the aforementioned plates are provided with the two opposite edges bonded to the plates. One group of fins are longitudinally positioned in spaced relation between the bottom and intermediate plates to form first fluid passageways. Another group of fins are longitudinally posiicc tioned in spaced relation between the intermediate and top plates to form second fluid passageways. The two fin groups form fluid passageways with adjacent ns spaced apart at density of 30-80 fins per inch transverse to the direction of fluid flow. The plates are preferably parallel to each other, and the fins preferably aligned normal to the plates.

The fins are between about 0.1 and 2 inches long, and longitudinally separated by gaps between about 0.03 and 0.25 inch long. The second fluid passageway gaps are superimposed over the first fluid passageway gaps. In a preferred embodiment, means are provided extending transverse the fins for transverse alignment of the gaps.

This heat exchanger is significantly more efficient than prior art plate-fin type constructions for several reasons. The fluid passageways are extremely narrow, so that the apparatus may be operated within the laminar flow region, i.e. with Reynolds numbers below about 2000. Under these circumstances the overall heat transfer coefficient is inversely proportional to the width of the passageways. For example, reducing the passage width by half doubles the heat transfer coefficient. In view of such high coefllcients, the length of the fluid passageway required to effect a required heat transfer may be greatly reduced. Shorter passageways in turn reduce the pressure drops of the fluids flowing therethrough. Lower pressure drops for a required heat exchange overcomes a severe limitation on improving the efficiency of prior art plate-fin type heat exchangers as previously discussed.

One of the reasons why the prior art has not employed plate-fin heat exchangers with the high fin densities characteristic of this invention is the excessively high longitudinal heat conduction to be expected. Such conduction occurs both through the metal plates separating the fins, and through the fins themselves. As a result, it would be expected that high fin density exchangers of short length would be no more efficient than conventional low fin density turbulent flow units. However, the problem of longitudinal heat conduction has been solved by two essential features of this invention. The first feature is the employment of relatively low thermal conductivity metal such as stainless steel for the plate material and relatively high thermal conductivity metal such as copper for the fin material. The second feature is very short fins (0.1-2 inches long) longitudinally positioned in spaced relation with adjacent fins longitudinally separated by gaps 0.03-0.25 inch long. The second fluid passageway gaps are superimposed over the first fluid passageway gaps. In this manner the heat is forced to flow from the warmer fluid in one passageway transversely through the high conductive fin to the low conductive separating plate. It then flows transversely through the plate and through the high conductive fin to the colder fluid in the second passageway with minimal longitudinal conduction.

Another aspect of this invention relates to an improved method for fabricating a plate-iin type heat exchanger. Metal strips are provided and their longitudinal edges folded over such that outer surfaces are more than with respect to the intermediate section between such edges. The folded metal strips are then cut in the transverse direction to form fins 0.1-2 inches long. Next the fins are stacked with longitudinal surfaces of adjacent fins in contiguous end-to-end association, and bonded together to form a multiplicity of stacks.

At least three at plates are provided and the n stacks are aligned between a bottom and intermediate plate, and between the intermediate and a top plate. The alignment is a sandwich configuration with tin longitudinal edges contiguously associated with the plates and having gaps 0.030.25 inch long separating ends of adjacent fin stacks in the longitudinal direction. In this manner first fluid passageways are formed between the bottom and intermediate plates, and second fluid passageways between the intermediate and top plates with the latters gaps superimposed over the first fluid passageway gaps. The n stackplate sandwich assembly is brazed together to bond the n edges to the plates. To complete the heat exchanger, iiuid inlet and outlet headers are installed against the outer edges of the stacked fins at opposite ends of the first and second fluid passageways.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a fragmentary isometric view of a heat exchanger section constructed according to one embodiment of the invention.

FIG. 2 is a cross-section view taken in elevation of the FIG. 1 heat exchanger along line 2 2.

FIG. 3 is a fragmentary isometric View of a heat exchanger section constructed according to another embodiment of the invention.

FIG. 4 is a cross-section view taken in elevation of the FIG. 3 heat exchanger along line 4 4.

FIG. 5 is a graph showing the heat transfer effectiveness versus fins per inch for this and prior art plate-1in type heat exchangers.

FIG. 6 is an isometric view of a complete heat exchanger constructed according to this invention, and

FIG. 7 is a plan view of a louvered 1in.

DESCRIPTION OF THE PREFERRED EMBODIMENTS Metals most useful as the fins and plates of this heat exchanger are listed in Table I, along with their thermal conductivities and certain satisfactory combinations of fins and plates having thermal conductivity ratios of at least 4.

TABLE I Thermal Conductivity,

]3.t.u./ft.2 hr. F. atw

Fin Material 150 R. 530 R.

Copper 310 226 Aluminum 80 95 Steel 35 38 Plate Material:

Stainless steel 5 8 70% copper, 30% nickel (Cuproniekel) 10 17 80% nickel, 14% chromium, 6% iron (Ineonel) 7 9 Thermal Fin/Plate Conductivity Thermal 1?.t.u./ft.2 hr. Conductivity F. a Ratios Fin/Plate Combinations 150 R. 530 R. 150 R. 530 R.

Copper/Stainless steel 310/5 226/8 62 28 CopperCupi-onickel 310/ 10 226/17 31 13 Aluminium/Stainless steel 80/7 J5/8 11 12 Stoel/Stainless steel 35/5 38/8 7 5 It will be apparent from Table I that the thermal conductivity values and ratios vary somewhat under operating conditions, depending upon the temperature of the heat exchanger. This variance is particularly noticeable if the heat exchanger is operated at cryogenic temperatures, e.g. about 150 R. Copper is the preferred 1in material and stainless steel the preferred plate material, so that the iin/plate thermal conductivity ratio between ambient and cryogenic temperature for the copper and stainless steel combination is 28-62. Copper is preferred because of its ease in forming, adequate strength, high thermal conductivity, and ease in metal bonding. Stainless steel is the prefered plate material beallse 0f its high 4 strength and relatively low thermal conductivity. Another advantage of these particular metals is that they may be bonded together by lluxless brazing procedures, thereby providing a good thermal bond but avoiding possible plugging of the narrow flow passages with welding ux material.

The fins should be at least about 0.1 inch long as shorter members would unduly increase complexity of fabrication with only marginal reduction in longitudinal heat conduction. However the lins should not be longer than about 2 inches, as the longitudinal heat conduction would become sufficient to limit the heat exchanger effectiveness. A lin length of 0.2-1 inch is preferred.

Because the closely spaced iins should be uniformly shaped and spaced parallel from each other to provide for uniform tiow within each passage and equal fluid flow distribution through the many narrow passages, the individual fins are preferably self-supporting and reasonably rigid. Accordingly, the fin material thickness and/ or stiffness should be adequate to provide a dimensionally stable iin which can be reliably brazed to the separating plates. However if the fins are too thick their spacing as determined by the folded back edges (discussed hereinafter) would be excessive, i.e. less than 30 fins/inch iin density. Fin thicknesses of 0.003-0015 inch represent a satisfactory balance of these considerations.

With respect to fin height, relatively low fins increase manufacturing complexity, while the probability of fin deformation and consequent flow maldistribution increases with large fin heights. A fin height range of 0.1-1.0 inch is satisfactory while heights of between about 0.3 and 0.6 inch are preferred.

Heat exchangers having iin densities less than about 30 ns/ inch do not usually operate eifectipely in the laminar flow ration where low pressure drop performance is obtained. Accordingly, the n density must be at least at this level to realize the high heat transfer performance of this invention. Densities above about fins/inch inordinately increase the manufacturing complexity as cornpared to improved heat transfer performance. This is because of the increased precision required in stamping out the individual fins, alignment between the separating plates and brazing the fin-plate assembly to form a heat exchanger core.

The separator plates should be sufficiently thick to provide adequate strength for withstanding the uid process pressures and also have suflicient structural rigidity for stacking during manufacturing. IOn the other hand, unnecessarily thick plates would undesirably increase the longitudinal heat conduction of the heat exchanger. Plate thicknesses of between about 0.02 and 0.04 inch have been found most satisfactory.

FIGS. 1 and 2 illustrate one embodiment comprising spaced fiat fins 11 longitudinally aligned in parallel rows a, b, c, d, e, f, and g. The upper side of bottom plate 12 is metal bonded to the lower edges of one series of fin rows whereas the lower side of intermediate plate 13 is metal ybonded to the upper edges of the same series of fin rows to form rst fluid passageways `14. The upper side of intermediate plate 13 is metal bonded to the lower edges of a second series of iin rows whereas the lower side of top plate 15 is metal bonded to the same series of iin rows to form second uid passageways 16. Although only two passageways 14 and 16 are illustrated for simplicity, it will be appreciated that additional passageways may be provided to process any desired throughput or number of heat exchanging fluid streams. These additional passageways would be superimposed over first pssageways 14 or secured beneath second passageways 1 It will be apparent from FIGS. 1 and 2 that means may be needed to transversely space the fins in adjacent longitudinal rows ag. One satisfactory spacer comprises strips positioned between adjacent fins, and of appropriate width to provide the desired tin density. These strips may be removed after the lin edges are bonded to the plates, or may be permanently installed.

In one embodiment, the rows ,of tins 11 are transversely spaced from each other by curling or folding over their respective opposite edges 17 adjacent to the plates 12, 13 and y15 so that the outer surface of each edge fold abuts an adjacent fin row. The angle and radius of the curled or folded edge 17 is selected to provide the desired fin spacing. The fold-over angle should exceed 90 and preferably be about 180 so that essentially a line contact is formed between n edges 17 and plates 12, 13 and 15. Such folded edges also improve the metal bonding operation because little brazing material 18 is needed to bond the n to the plate at the line of contact due to the molten brazing materials capillary action. Minimum brazing material is desirable to minimize longitudinal heat conduction in the heat exchanger and to avoid the possibility of plugging the narrow flow passages.

It should be appreciated that although the FIGS. 1-2 embodiment shows ns 11 aligned end-to-end in the longitudinal direction, this is preferred for ease of fabrication but not essential. The end-to-end fins may alternatively be assembled in offset through parallel relation so that rows a-g would be non-existant.

Because of its high heat transfer effectiveness, the instant heat exchanger is quite short thereby greatly increasing the need for limiting the longitudinal heat transfer. For example, the heat exchangers of this invention are characterized by high longitudinal temperature gradients above about 30 F. per foot of length. In view of this unique characteristic, it is necessary to completely interrupt or separate adjacent fins 11 of each row a-g in the longitudinal direction. As illustrated in FIGS. 1 and 2, such adjacent fins 1l are longitudinally positioned with a separating gap or space 19 between ends.

For effective interruption of longitudinal heat transfer along fin rows a-g, it has been found necessary to transversely align the gaps 19 in the adjacent rows with the second fluid passageway gaps superimposed over the rst uid passageway gaps. Otherwise, heat could be shortcicuited across intermediate plate 13 between overlapping tins 11 and around gaps 19. Also it is necessary to substantially exclude all brazing or solder material from gaps 19. Such gaps should be between about 0.03 and 0.25 inch long. In general shorter gaps are preferred with shorter fins and conversely the longer gaps are most advantageous with longer fins. If the gaps are less than 0.03 inch the longitudinal heat transfer is not sufficiently interrupted, whereas gaps longer than 0.25 inch substantially reduce the n surface area available for lateral heat transfer through the plates. Also, the latter results in excessively long heat exchanger constructions Without further reducing the longitudinal heat transfer by an amount justifying the additional heat exchanger length. Fin gaps of 0.03-0.l inch are preferred.

Means extending transverse to the ns through gaps 19 in adjacent iin rows a-g are employed to insure the required transverse alignment. Wires or rods 20 are well suited for this purpose, their cylindrical cross section affording only a point-to-point contact with the fins thereby minimizing longitudinal heat transfer. For the same reason they are preferably positioned only by frictional contact with the n ends `without metal bonding thereto. Wires 20 are preferably bent into a wave pattern with projecting sections on at least one side contacting plate 12, 13 or so as to be self-supporting within gaps 19. Instead of extending through gaps 19, wires may be passed through holes drilled in fins I11 for transverse alignment thereof. Moreover the wires 20 could be used only during the heat exchanger fabrication and removed once the fins have been positioned with the desired transverse alignment of gaps. Metal strips could be used instead of wires 20.

FIGS. 3 and 4 illustrate another heat exchanger embodiment of the invention similar to the FIS. 1-2 construction, but also differing in certain particulars to be described in detail. Corresponding elements have been assigned the same numeral designations for clarity.

Longitudinal heat conduction along fins 11 may be further reduced by cutting a multiplicity of narrow spaced crosswise slots 21 in the fins normal to the iin edges. Such slots 21 are most effective with relatively long ns, e.g. between 1 and 2 inches long. They serve to reduce the width of the solid conductive longitudinal heat transfer path along the ns and thereby decrease the amount of heat transferred from one end of a fin to the other end. Slots 21 are also useful in the individual parallel fins illustrated by FIGS. 1 and 2.

Another features of FIGS. 3 and 4 is the use of a single continuous sheet to form a plurality of the fins of each passageway 14 and 16. In the FIGS. 1-2 embodiment each fin is separate and distinct from the other fins, and must be separately bonded to the plates 12, 13 and 15. In the FIGS. 3-4 embodiment a single sheet is folded or corrugated to form the desired number of parallel spaced tins. One advantage of this arrangement is that positioning and bonding of the fins to the plates is simplified. Due to the continuous nature of the fin assembly it is not essential to employ separate means for transversely spacing the fins, as for example the folded edges 17 of FIGS. l-2. A disadvantage of the corrugated fins is that longitudinal heat transfer is increased somewhat due to the solid metal folds joining adjacent ns.

The gaps 19 between adjacent fins 11 of each fin row are transversely aligned by deformities 20a raised from plates 12, 13 and 15 extending transverse the plates and normal to the fin rows. These raised deformities are spaced from each other by an appropriate distance so that the adjacent fin rows are stopped by the uphill and downhill sides of a given deformity 20. Alternatively the gaps 19 could be formed by positioning strips in spaced relation across the plates 12, 13 and 15 to stop adjacent fins.

The heat transfer effectiveness of various plate-fin heat exchangers of this invention in the laminar flow region is compared in FIG. 5 with prior art plate-fin exchangers in the turbulent ow region. FIG. 5 shows the heat transfer effectiveness (and thus economy) as the ordinate, plotted against the n desnsity or number of tins per inch as the abscissa. Heat transfer effectiveness is a wellknown measure of performance of a particular heat exchanger as compared with the ideal. In particular, effectiveness is the ratio of the actual heat transferred to a colder fluid to the heat which would be transferable in a heat exchanger having infinite area and zero nal temperature difference. In a heat exchanger having rst and second passageways for balanced flows of two identical fluids, the heat transfer effectiveness is the ratio of the temperature rise of the incoming fluid being warmed (or temperature drop of the uid being cooled) to the temperature difference between the incoming warm and cold fluids, and is always less than unity. The point and curves are based on the same total heat transfer area, a constant pressure drop across the exchanger, 0.005 inch n thickness, 0.420 inch fin height, and 0.02 inch plate thickness. The basis for comparison is a heat transfer effectiveness of 0.96 (point AA) for the conventional turbulent flow units having 18 tins/inch and lacking the fin gaps and using aluminum as the construction material for both the fins and plates.

Curve A shows the theoretical relationship for simple plate-1in construction and neglecting the detrimental effect of longitudinal heat conduction on thermal performance. It can be seen that with laminar iiow, the higher theoretical performance is achieved by increasing the n density. Curve B shows the effect of longitudinal heat conduction on iin density by virtue of using a high thermal conductivity metal, copper, for both fins and plates with laminar flow. It will be seen that the laminar flow unit about matches the turbulent flow unit criteria of 0.96 effectiveness at lin densities of about 4050 fins per inch, and decreases rapidly at higher n densities. This decrease occurs because higher fin densities require progressively shorter heat exchangers to satisfy the constant AP, thereby greatly increasing the effect of the longitudinal heat conduction. Such construction would not represent an improvement over the turbulent flow plate-fin heat exchangers.

Curve C shows the heat transfer performance of a heat exchanger of this invention with copper fins and stainless steel plates. This heat exchanger has fins 1 inch long and spaced with 0.06 inch gaps. The curve indicates an optimum performance of about 97.5 heat transfer effectiveness at about 55-60 fins/inch. Curve D shows the further improvement attainable by shortening the fins to 0.25 inch long also spaced with `0.06 inch gaps. This increases the number of longitudinal heat transfer interruptions, and provides about 98.1 heat transfer effectiveness at about 70-90 fins/ inch.

Another aspect of this invention relates to an improved method for fabricating plate-lin type heat exchangers including those previously described. In this method, metal strips are provided and formed into fins by folding over the longitudinal edges such that the outer surfaces are more than 90 and preferably 180 with respect to the intermediate section between the edges. This step may for example be performed by passing the strip material through a die. The folded strips are then cut in the transverse direction to form ns between about 0.1 and 2 inches long, and would appear as fins 11 in FIGS. 1-2.

Next the individual ns are stacked with longitudinal surfaces of adjacent fins in contigous end-tO-end association, and the stacked fins are attached together to form a multiplicity of stacks. The attachment need only be strong enough to hold each stack together for subsequent handling and alignment between plates for the nal brazing step. Accordingly the attachment may be provided by a suitable plastic adhesive binder or alternatively the fins may be joined by head welding along the opposite corners of each stack. Again referring to FIG. 1, a stack may comprise fins a, b, c, d, e, f and g separated by spaces 16.

Next one group of the fln stacks is aligned between a bottom and intermediate flat plate, and another group of fin stacks is aligned between the intermediate and a top flat plate in a sandwich configuration having fin longitudinal edges contiguously associated with the plates. The alignment is such that gaps 0.03-0.25 inch long separate ends of adjacent n stacks in the longitudinal direction.

In this manner first fluid passageways are formed between the bottom and intermediate plates and second fluid passageways between the intermediate and top plates with the latters gaps superimposed over the first fluid passageway gaps.

For metal bonding the n stacks to the flat plates, a suitable brazing foil preferably in the form of thin strips is positioned between the two components in the longitudinal direction, from one end to the other end of each stack. The brazing foil preferably should not extend into the gaps separating adjacent iin stacks, as the metal would at least partially fill the gaps on melting and increase longitudinal heat conduction. The brazing foil composition is of course selected to provide a good bond between the metals selected for the ns and plates, although this bond need not be uid-tight. For example, in the case of copper fins and stainless steel plates, an alloy comprising 5 wgt. percent tin, 17% zinc, 22% copper and 56% silver has been found suitable in the form of 0.004 inch thick foil.

The iln stack-plate sandwich assembly is brazed in the well-known manner to bond the fln edges to the plates. The fluxless furnace brazing technique is preferred; the assembly may for example be placed in a closeable vessel or retort and the latter positioned in the furnace. If a plastic binder is used to temporarily hold the sandwich together the retort is preferably initially heated to 800- 900 F. in a reducing atmosphere such as hydrogen gas to dissipate the binder. The temperature is then increased to at least 1400 F. still in a hydrogen atmosphere and as rapidly as possible. When the desired maximum temperature is reached the heating is terminated and the retort cooled to about l000 F. before opening. When the latter has cooled to about 300 F. the retort is purged with an inert gas such as nitrogen to remove the residual hydrogen before opening to the atmosphere.

Side plates, manifold and header assemblies are then installe-d as will be understood by those skilled in the heat exchanger art.

A suitable completed heat exchanger constructed according to this invention is illustrated in FIG. 6 with six fluid passageways and several fin stacks in each passageway. For example, the uppermost passageway 22 comprises several fn stacks 23 positioned in longitudinal endto-end relation with intervening gaps 19. Side plates 24 are preferably metal bonded to the edges of adjacent plates for rigidity and to prevent fluid leakage between adjacent iluid passageways.

To uniformly distribute the fluid ilow between adjacent fins of each passageway, orifice plates 25 may be transversely positioned at least at the inlet end and bonded to the side plates 24. Alternatively lin-type distributors could be employed although they tend to increase the heat exchanger overall length.

The FIG. 6 heat exchanger is intended for processing of two separate iluids with one fluid in rst passageways 14` and the other fluid in intervening second passageways 16. For manifolding, end closure plates 26 and 27 are transversely positioned in the ends of passageways 14 and 16 respectively. The plates extend only part way across the passageways and are coextensive with ilrst and second passageway hea-ders 28 and 29, respectively. These headers are only shown at one end of the heat exchanger to permit illustration of the orifice-manifold closure plate asembly at the opposite end. In a completed exchanger headers 28 and 29 are required at both ends.

The invention will be more clearly understood by reference to the following examples.

EXAMPLE I A small scale prototype plate-iin heat exchanger was fabricated in the aforedescribed manner with the tins in longitudinal rows, but without specic means for transverse alignment of the gaps. The copper fins were 0.005 inch thick x 0.5 inch wide x 1.5 inches long with the longitudinal edges curled over 180 for spa-cing. These folds were 0.020 inch long. l

The llat plates were 0.0175 inch thick stainless steel with a 1 mil thick copper plating on both sides for easier bonding of the ns. It should be recognized that this invention contemplates the use of fins and plates composed of dissimilar metal layers in which the fins have an average thermal conductivity at least four times the average thermal conductivity of the plates. In this example the stainless steel contributes a thermal conductivity of whereas the copper adds so that the average thermal conductivity for the plate is :31.2. The thermal conductivity ratio of the fins to the composite plates was '9 inches wide X 10 inches long, and included 1200 fins. It was tested for balanced flow conditions using heated air, that is, the air was heated in the rst passageway and countercurrently cooled in the second passageway. Typical data is summarized below in Table II.

TABLE II Gas Temperatures, F. Heat Cold Stream Warm Stream Trans- Gas Pressure, Flow fer Effecp.s.i.g. c.f.h. In Out In Out tiveness* *Based on temperature change of the Warmer stream divided by temperature diierenee between two incoming streams.

It will be noted that the heat transfer effectiveness btained for this unit was considerably below the 0.96-1- desired. These relatively poor results were due to only about 65% of the fins being well bonded to the flat plates, thus reducing heat transfer capacity proportionately.

EXAMPLE 2 Another small scale prototype heat exchanger was constructed in the aforedescribed manner with copper fins 0.005 inch thick X 0.5 inch wide x 1.5 inches long spaced at 63 ns/ inch density in longitudinal rows and 0.09 inch gaps. The ilat plates were 0.020 inch cupronickel and the fins were brazed thereto using 0.004 inch thick foil of an alloy comprising 5% tin, 17% zinc, 22% copper, and 56% silver. For transverse alignment of the gaps in the iin longitudinal rows, 0.06 inch diameter holes were punched near each end of the ns and 0.055 inch diameter stainless steel rods were inserte-d through these holes. Again the fins were transversely spaced from each other by curling over the longitudinal edges 180.

1 0 EXAMPLE In A larger heat exchanger of the type schematically illustrated in FIG. 6 was constructed as previously described using copper iins 0.005 inch thick X 0.46 inch high x 1.5 inches long. As illustrated in FIG. 7 each of these ns 11 contained three equally spaced lateral slots 21 (see FIGS. 344 for elevation View), and the two opposite n edges 17 were folded back 180 with folds 0.020 inch long (see FIGS. 1-2 for elevation View). The longitudinal sections 39 between each end and the adjacent lateral slot, and between the slots themselves were bent away from the fin longitudinal axis a-a, i.e. louvered. The purpose of these eccentric sections was to open the slots more widely for improved gas circulation across the iin rows and reduced longitudinal heat conduction. With shorter tins, e.g. 0.5 inch, the louverng would probably not be advantageous. As in Example II, 0.06 diameter holes 31 were punched near each end of the tins for insertion of 0.55 inch diameter stainless steel rods for gap alignment purposes.

The fins were positioned between 0.020 inch thick cupronickel plates in longitudinal rows with 0.06 inch gaps between adjacent fins in each row and aligned by the stainless steel rods. The as-assembled unit with 63 fins per inch and without headers was 3.25 inches thick x 5 inches wide X 36 inches long.

After placement of 0.003 inch thick strips of the brazing foil along each side of the ns in the longitudinal direction, the assembled unit was furnace brazed in a hydrogen atmosphere using the aforedescribed procedure. The orice plates, end closure plates and headers were then welded in position to complete the exchanger.

The FIG. 6 heat exchanger was also tested in the air heating-cooling system. Test series were run at warm end inlet temperatures of 160 F., 250 `F. and 340 F. Selected data obtained during the 250 F. series of tests are listed below in Table IV. This data is also representa-v tive of the 160 F. and 340 F. runs.

TABLE IV Gas Temperatures, F.

Heat Transfer Cold Side Warm Side Effectiveness Gas Pressure Flow, Reynolds p.s.i.g. lb./hr. In Out In Out No. 1l Avg.2

1 Based on temperature change of the warmer stream. 2 Based on average of temperature changes for Warmer and colder streams.

The iins were assembled in two uid passageways as illustrated in FIGS. l and 2 to form a heat exchanger approximately 1 inch thick x 2 inches wide X l0 inches long, using a total of 1200 tins.

The heat exchanger was then tested in the same manner as the Example 1 unit. Typical data is summarized below in Table III for warm end inlet temperatures of about 150 F.

TABLE III Gas Temperatures, F.

Cold Side Warm Side Heat Trans- Gas pressure, Flow, Reynolds fer Etecp.s.i.g. lb./l1r. In Out In Out No. trveness* Based on temperature change of Warnier stream compared to temperature difference between two incoming streams.

It is believed that higher values for heat transfer effecheat exchanger is suitable for cooling feed gas containtiveness would have been obtained using shorter fins, e.g. ing hydrocarbons, carbon oxides and hydrogen at about 0.5 inch instead of 1.5 inch.

300 p.s.i.a. by hydrogen gas at about 20 p.s.i.a. for par- 11 tial liquefaction. The heat exchanger flow directions in adjacent uid passageways may be countercurrent, cocurrent or even at right angles to each other, as will be understood by those skilled in the heat exchange art.

Although certain embodiments have been described in detail it will be appreciated that other embodiments are contemplated along with modifications of the disclosed features, as being within the scope of the invention.

What is claimed is:

1. A plate-fin type heat exchanger comprising:

(a) at least three flat plates in spaced superimposed relation with the intermediate plate over the bottom plate, and the top plate over the intermediate plate, said plates being formed of low thermal conductivity metal;

(b) a multiplicity of fiat fins formed of metal having high thermal conductivity at least 4 times that of said plates (a) with opposite edges bonded thereto, said fins being longitudinally positioned in spaced relation between said bottom and intermediate plates to form first fluid passageways, and between said intermediate and top plates to form second fiuid passageways with adjacent fins spaced at density of 30-80 fins per inch transverse to the direction of fluid flow, the fins being 0.1-1.0 inch high, 0.1-2 inches long and longitudinally separated by gaps 0.03-0.25 inch long with the second iiuid passageway gaps superimposed over the first uid passageway gaps.

2. A plate-1in type heat exchanger comprising:

(a) at least three fiat plates in spaced superimposed relation with the intermediate plate over the bottom plate, and the top plate over the intermediate plate, said plates being formed of low thermal conductivity metal;

(b) a multiplicity of flat ns formed of metal having high thermal conductivity at least 4 times that of said plates (a) with opposite edges bonded thereto, said fins being longitudinally positioned in spaced relation between said bottom and intermediate plates to form first fluid passageways, and between said intermediate and top plates to form second fluid passageways with adjacent fins spaced at density of 30-80 ns per inch transverse to the direction of uid ow, the fins being 0.1-1.0 inch high, 0.1-2 inches long and longitudinally separated by gaps 0.03-0.25 inch long with the second fluid passageway gaps superimposed over the first fluid passageway gaps, and

(c) means extending transverse said fins for transverse alignment of said gaps in each of said rst and second iiuid passageways.

3. A plate-fin type heat exchanger according to claim 2 in which said means (c) are positioned in said gaps.

4. A plate-fin type heat exchanger comprising:

(a) at least three flat plates in spaced superimposed relation with the intermediate plate over the bottom plate, and the top plate over the intermediate plate, said plates being formed of low thermal conductivity metal;

('b) a multiplicity of fiat fins formed of metal having high thermal conductivity at least 4 times that of said plates (a) with opposite edges bonded thereto, said fins being longitudinally aligned in parallel spaced rows between said bottom and intermediate plates to form first fluid passageways, and between said intermediate and top plates to form second fluid passageways with adjacent fins spaced at density of 30-80 fins per inch transverse to the direction of fluid flow, the fins being 0.1-1.0 inch high, 0.1-2 inches long and longitudinally separated by gaps 0.03-0.25 inch long with the secondfluid passageway gaps superimposed over the first fiuid passageway gaps; and

(c) means extending transverse said ns through said gaps in adjacent fin rows for transverse alignment of said gaps in each of said first and second fluid passageways.

5. A plate-iin type heat exchanger according to claim 1 in which said plates are formed of stainless steel and said fins are formed of copper.

6. A plate-fin type heat exchanger according to claim 1 in which said fins are 0.2-1 inch long and said gaps are 0.03-0.10 inch long.

7. A plate-tin type heat exchanger according to claim 1 in which said edges of said fins are folded more than to form outer surfaces longitudinally bonded to said plates and abutting an adjacent fin for transverse spacing therefrom.

8. A plate-1in type heat exchanger according to claim 1 in which edges of said fins are folded 180 to form outer surfaces having fold edges longitudinally bonded to said plates and abutting an adjacent fin for transverse spacing therefrom.

9. A plate-fin type heat exchanger according to claim 1 with a multiplicity of narrow crosswise slots in the fins normal to the fin edges and spaced from each other.

10. A plate-fin type heat exchanger according to claim 1 in which a multiplicity of corrugated sheets transversely aligned between said plates and longitudinally spaced from each other comprise said fins.

11. A plate-fin type heat exchanger according to claim 2 in which a multiplicity of spaced rods extending across the heat exchanger width comprise the gap transverse alignment means.

12. A plate-fin type heat exchanger comprising:

(a) at least three fiatv stainless steel plates in spaced superimposed relation with the intermediate plate over the 'bottom plate and the top plate over said intermediate plate;

(b) a multiplicity of flat copper fins with opposite edges folded to form outer surfaces 1ongitudi nally ybonded to said plates and abutting transversely spaced fins to align said fins in parallel spaced rows between said bottom and intermediate plates to form first fluid passageways, and between said intermediate and tp plates to form second fluid passageways, with adjacent fins spaced at density of 30-80 fins per inch transverse to the direction of uid iiow, the tins being 0.1-1.0 inch high, 0.1-2 inches long and longitudinally separated -by gaps 0.03-0.25 inch long with the second fluid passageway gaps superimposed over the first fluid passageway gaps; and

(c) a multiplicity of spaced rods each extending across the heat exchanger width through adjacent gaps in adjacent fin rows for transverse alignment of said gaps.

13. A plate-iin type heat exchanger according to claim 1 with a multiplicity of narrow crosswise slots in the fins normal to the fin edges and spaced from each other, the fin sections between said slots ybeing louvered.

, References Cited UNITED STATES PATENTS 1,899,080 2/1933 Dulgliesch 16S-166 2,566,310 9/1951 Burns et al. 165-167 X 2,601,973 7/1952 Jensen 16S-166 3,313,343 4/1967 Ware et al. 16S-166 2,663,550 12/1953 Hammond et al 165-166 X 2,870,998 l/l959 Woolard 16S-166 X 3,079,994 3/1963 Kuehl 16S-166 3,282,334 11/1966 Stahlheber 16S-166 2,549,466 4/1951 Huheisel 165-166 X 2,875,986 3/1959 Holm 16S-166 X 2,952,445 9/1960 Ladd 16S-166 3,148,442 9/1964 Gier 16S-166 X ROBERT A. OLEARY, Primary Examiner T. W. STREULE, JR., Assistant Examiner U.S. Cl. X.R. 16S-180 

